Method and system for controlling a transfer case clutch to avoid wheel slip

ABSTRACT

In a motor vehicle driveline including a transfer case whose output is continually connected to a first output, a clutch, operating partially engaged, responds to a control signal to change the degree of clutch engagement, whereby a second output is connected driveably to the first output. A digital computer, repetitively executing a computer readable program code algorithm for operating the clutch partially engaged, continually selects a desired magnitude of clutch engagement with reference to functions indexed by vehicle speed and either engine throttle position or engine throttle rate. The computer repetitively updates at frequent intervals the desired degree of clutch engagement, and issues a command clutch duty cycle to a solenoid-controlled valve, which signal changes the degree of clutch engagement in response to the command signal.

BACKGROUND OF INVENTION

1. Field of the Invention

This invention relates to the field of transfer cases for motorvehicles; more particularly, it pertains to control of a transfer caseclutch.

2. Description of the Prior Art

A transfer case is a device located in a motor vehicle driveline betweenthe output of a geared power transmission and front and rear driveshaftsfor transmitting power to the wheels. A transfer case may include aplanetary gear set that produces both a high range, in which thetransfer case output is driven at the same speed as the input, and a lowrange, in which the output is driven slower than its input speed. The4×2 and 4×4 states of the transfer case are usually selected manually bythe vehicle operator by operating a lever or switch. A first position ofthe lever causes a range selection device in the transfer case to directpower from the transmission output to a rear drive axle, the 4×2-drivemode. A second position of the lever causes the transfer case to directpower to both a front drive axle and a rear drive axle, the 4×4-drivemode.

Certain transfer case control systems either fully engage or fullydisengage the secondary driveshaft and the power source. A transfer casecontrol system for all wheel drive operation transmits power continuallyand variably to the front and rear driveshafts. Various techniques areavailable for establishing the torque split or portion of the enginetorque that is transmitted to the front and rear wheels. For example, acenter differential mechanism continually divides torque at a fixedratio between the front and rear wheels perhaps 35% of torque to thefront wheels and 65% to the rear wheels. But a center differentialmechanism provides no variation of the torque division as needed toimprove vehicle handling under certain drive conditions. A centerdifferential usually includes a planetary gearset having an input, suchas a sun gear driven by the transmission output, a first output such asa carrier connected to the rear driveshaft, and a second output such asring gear connected to the front driveshaft.

A viscous coupling, located in parallel with the front and reardriveshafts, or the first and second outputs of a center differential,operates to mutually connect or couple the driveshafts in proportion tothe speed difference between them. It produces this effect by shearing aviscous fluid located between closely spaced plates, one set of platesdriven by the front driveshaft and a second set of plates driven by therear driveshaft. Variations in the speed difference of the sets ofplates increase the magnitude of the forces tending to maintain theplates at the same speed. The coupling dissipates a portion of theoutput power in the process of synchronizing the speeds of the front andrear driveshafts.

A hydraulically actuated clutch continually driveably connected to aprimary driveshaft can be used to transmit a variable magnitude oftorque to a secondary driveshaft. The magnitude of torque transmitted tothe secondary shaft is controlled electronically to improve vehiclehandling characteristics under certain drive conditions. Engine throttleposition and the time rate of change of throttle position over the fullrange of vehicle speeds present a basis for improving vehicle handlingthrough control of the clutch.

For example, when vehicle speed is high and the transmission isoperating in a high gear ratio, if the vehicle operator then tips intothe throttle, a low magnitude of torque at the front wheels is requiredto prevent or preempt wheel slip. However, when vehicle speed is low andthe transmission is operating in a low gear ratio, if the vehicleoperator then tips into the throttle, a relatively high magnitude oftorque at the front wheels is required to prevent wheel slip.

In another example, when throttle rate is high, either due to a throttletip-in or back-out, a relatively high magnitude of torque at the frontwheels is required to prevent or preempt wheel slip. However, when thethrottle rate is low, a relatively low magnitude of torque at the frontwheels is required to prevent wheel slip due to the inertia of theengine and transmission. At low vehicle speed with the transmissionoperating in a low gear ratio, a relatively high magnitude of torque atthe front wheels is required to prevent wheel slip.

There is a need, therefore, for a method, system and apparatus forcontrolling predictably and reliably a transfer case clutch toaccommodate the particular driveline related consequences andrequirements of engine throttle position, and the time rate of change ofthrottle position over the full range of vehicle speed to improvevehicle handling, such as the avoidance of wheel slip.

SUMMARY OF INVENTION

According to this invention, in a motor vehicle driveline that includesa transfer case whose output is continually connected to a first output,a clutch, operating partially engaged, responds to a control signal tochange the degree of clutch engagement. A digital computer, repetitivelyexecuting a computer readable program code algorithm for operating theclutch partially engaged, selects a desired magnitude of clutchengagement with reference to functions indexed by vehicle speed andeither engine throttle position or engine throttle rate. The computerrepetitively updates at frequent intervals the desired degree of clutchengagement, and issues a command clutch duty cycle to asolenoid-controlled valve, which signal changes the degree of clutchengagement in response to the command signal.

The control method and system of this invention accommodates theparticular driveline related requirements of engine throttle positionchanges, tip-ins and back-outs, and the time rate of change of throttleposition over the full range of vehicle speed to improve vehiclehandling, particularly the avoidance of wheel slip.

In realizing these and other advantages, a method, according to thisinvention, for controlling a clutch that driveably connects an input andan output with varying degrees of clutch engagement in a vehicle driveline having an engine controlled by a throttle position, includes thesteps of operating the clutch partially engaged, and determining thecurrent throttle position, throttle rate, and vehicle speed. Then both afirst desired clutch engagement corresponding to the current throttleposition and vehicle speed, and a second desired clutch engagementcorresponding to the current throttle rate and vehicle speed aredetermined. The degree of clutch engagement is changed to the greater ofthe first desired clutch engagement and the second desired clutchengagement.

Another embodiment of this invention contemplates a method forcontrolling, with the aid of a digital computer, a clutch that driveablyconnects an input and an output with varying degrees of clutchengagement, the clutch operating in a vehicle driveline that includes anengine controlled by throttle position. The method includes first thestep of inputting to and executing in the computer a computer readableprogram code algorithm for operating the clutch partially engaged. Thecomputer is provided with a data base that includes at least a firstdesired clutch engagement that varies with the current throttle positionand vehicle speed, and a second desired clutch engagement that varieswith a current throttle rate and a current vehicle speed. Signals areprovided at frequent intervals to the computer representing the currentthrottle position, and vehicle speed. The computer repetitivelydetermines at frequent intervals during execution of the algorithm thecurrent throttle rate, the first desired clutch engagement correspondingto the current throttle position and vehicle speed, and the seconddesired clutch engagement corresponding to the current throttle rate andvehicle speed. Subsequently the computer issues a command clutch dutycycle signal representing the greater of the first desired clutchengagement and the second desired clutch engagement. This commandchanges the current degree of clutch engagement to the degree of clutchengagement represented by the command signal.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a top view of a motor vehicle driveline having a transmission,transfer case, and drive shafts extending to front wheels and rearwheels;

FIGS. 2A and 2B are left-hand and right-hand portions, respectively, ofa cross sectional side view of a transfer case and a portion of anautomatic transmission;

FIG. 3 is a schematic diagram of a system for controlling a transfercase clutch in a four-wheel drive vehicle;

FIG. 4 is a block diagram showing the steps for controlling a transfercase clutch;

FIG. 5 is a block diagram showing the steps for controlling a transfercase clutch in a handling mode condition; and

FIG. 6 is a block diagram showing the steps for controlling a transfercase clutch to protect against overheating.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

With reference now to the drawings and particularly to FIG. 1, thepowertrain of a motor vehicle, to which the present invention can beapplied, includes front and rear wheels 10, 12, a power transmission 14for producing multiple forward and reverse speed ratios driven by anengine (not shown), and a transfer case 16 for continuously driveablyconnecting the transmission output to a rear drive shaft 18. Thetransfer case 16 selectively connects the transmission output to boththe front drive shaft 20 and rear drive shaft 18 when a four-wheel drivemode of operation is selected, either manually or electronically. Shaft18 transmits power to a rear wheel differential mechanism 22, from whichpower is transmitted differentially to the rear wheels 12 through axleshafts 24, 26, which are contained within a differential housing. Thefront wheels are driveably connected to right-hand and left-hand axleshafts 32, 34, to which power is transmitted from the front drive shaft20 through a front differential mechanism 36.

It should be appreciated that the terms “front” and “rear” are usedherein for convenience purposes only, to refer to a secondary andprimary driveshafts, respectively. In alternate embodiments of theinvention, the front and rear driveshafts may be interchanged, e.g., thefront driveshaft may act as the primary driveshaft.

Referring now to FIGS. 2A and 2B, the output shaft 38 of the automatictransmission 14 extends through the transmission casing 37 into thecasing 16 of the transfer case. Shaft 38 is driveably connected througha spline 40 to the sun gear 42 of a simple planetary gear set, anepicyclic train 44. Sun gear 42 is in continuous meshing engagement witha set of planet pinions 52, which are supported for a rotation on stubshafts 54, each stub shaft supported at opposite axial ends on a carrier56. Each of the planet pinions 52 is in continuous meshing engagementwith the sun gear 42 and a ring gear 46. Carrier 56 is driveablyconnected through spline 57 to the output 58 of the transfer case, whichis adapted for connection to the rear driveshaft 18.

A high-low coupler 60 includes a hub 62, which is driveably connectedthrough a spline 59 and radial disc 63 to ring gear 46. Coupler 60includes a sleeve 64, formed on its inside surface with a system ofaxially directed spline teeth 66, engaged continuously with a system ofspline teeth 68 formed on the outer surface of the hub 62. The sleeve 64slides axially leftward and rightward on the hub. In FIG. 2, the coupler60 shown above the axis of output shaft 58 is a synchronizer; thecoupler shown below that axis is a dog clutch.

The teeth 66 of the sleeve 64 are engageable alternately with axiallydirected spline teeth 70 formed on a radially outer surface of a disc72, which is continually fixed against a rotation by its engagement at74 with teeth formed on the inner surface of the transfer case 16. Theteeth 66 of sleeve 64 are engageable also with a system of axiallydirected spline teeth 76 formed on a radially outer surface of a disc78.

Disc 78 is splined at 79 to carrier 56, which is splined at 57 to outputshaft 58. Spline 81 driveably connects shaft 58 to a drum 82, which isformed on its inner surface with axially directed spline teeth 84.Spacer plates 86 are driveably engaged with the spline 84 of drum 82.Friction discs 88, interposed between adjacent spacer plates 86, aredriveably engaged by spline teeth formed on the outer surface of an arm91, which extends axially from a drive belt sprocket wheel 92.

Located within drum 82 is a hydraulically actuated piston 94, whichmoves axially in response to the pressurized and vented state of ahydraulic cylinder 96 located between drum 82 and piston 94. Whencylinder 96 is pressurized, piston 94 moves rightward forcing the spacerplates 86 and friction discs 88 into mutual frictional engagement,thereby driveably connecting output 58 and sprocket wheel 92. Whencylinder 96 is vented, piston 94 is moved leftward to the position shownin FIG. 2 due to a force applied to the piston by a Belleville spring98, thereby driveably disconnecting output 58 and sprocket wheel 92. Inthis way, clutch 100 alternately driveably connects and disconnectsoutput 58 and sprocket wheel 92.

When clutch 100 is engaged, power is transmitted to the forward driveshaft 20 from the output shaft 58 by a drive belt 90, which iscontinually engaged with sprocket wheel 92. Bearings 95, 96 rotatablysupport sprocket wheel 94 on the transfer case 16, and forward driveshaft 20 is driveably connected through a spline 98 formed on the innersurface of the sprocket wheel 94. In this way, when clutch 100 isengaged, output shaft 80 transmits power both to the rear drive shaft18, which is connected by a universal joint to output shaft 80, and tothe forward drive shaft 20.

In operation, drive shaft 20 is driven alternately at the same speed asthat of the transmission output shaft 38, or shaft 20 is underdriven inrelation to the speed of shaft 38, in accordance with the position ofthe coupler sleeve 64.

Carrier 56 is continually driveably connected to output shaft 58 throughspline 57. Ring gear 46 is driveably connected to output shaft 58through the torque delivery path that includes disc 63, coupler hub 62,coupler sleeve 64, disc 78 and splines 79, 57. Therefore, when sleeve 64moves rightward to the position shown in FIG. 2, ring gear 46 andcarrier 56 are mutually driveably connected, and ring gear 46, carrier56 and output 58 are driven at the same speed as that of sun gear 42 andthe input 38. This is the high-speed range.

When sleeve 64 of coupler 60 is moved leftward to produce a driveconnection between disc 72 and coupler hub 62, ring gear 46 is fixedagainst rotation on the transfer case 16 through the torque path thatincludes disc 63, coupler hub 62, its sleeve 64 and disc 72. Thisprovides a torque reaction and causes carrier 56 and output 58 to beunderdriven in relation to the speed of sun gear 42 and shaft 38. Thiscreates a low-range drive connection between transmission output 38 andthe transfer case output 58.

Clutch 100 can be engaged regardless of the position of coupler sleeve64 so that power is transmitted by the drive belt mechanism, whichincludes sprocket wheels 92, 94 and drive belt 90. In this way, both theforward drive shaft 20 and rear drive shaft 18 are driven alternately inthe low-range and high-range, or only the rear drive shaft is driven inthe low-range and high-range.

The transmission output 38 is driven by a ring gear 160, which issecured through a park gear 162 to output shaft 38. The park gear andshaft 38 are supported on the transmission case 37 by a bearing 164. Theouter surface of the park gear is formed with teeth 166 separated byspaces adapted for engagement by a park mechanism.

Referring now to FIG. 3, there is shown a transfer case control systemdeployed in a four-wheel drive vehicle. Torque produced by the engine169 is transmitted to the transfer case 16 at multiple speed ratiosproduced by an automatic transmission 14 located in the torque pathbetween the engine 169 and transfer case 16. The rear drive shaft 18 iscontinually connected to the transfer case output 58, but clutch 100transmits a variable portion of that input torque to the frontdriveshaft 20 in accordance with the degree to which clutch 100 isengaged. Clutch slip, the difference in speed of the rear driveshaft 18and front driveshaft 20, is a measure of the degree of engagement ofclutch 100.

The transfer case control system 10 includes a microcontroller ordigital computer 170 operating under stored program control having acentral processing unit that includes an arithmetic logic unit;electronic memory 172 containing control algorithms, data, functions,scalar values and routines for producing data and informationrepresenting current operating conditions of the powertrain; andinput/output devices for acquiring, conditioning and transmittingsignals produced by data sensors and control devices that respond tocommand signals generated by the computer. Controller 170 furtherincludes continual data communication links among its CPU, ALU, I/O,memory, control algorithms and operational controls.

Controller 170 receives signals generated by sensors 144-148 and171-168, processes and uses the input signals to determine the amount oftorque to be transmitted to front and rear driveshafts 20, 18,respectively. Based upon this determination, controller 170 generates acommand signal to activate the clutch 100, thereby selectivelytransmitting torque to the front driveshaft 20 from the transfer caseoutput 58.

In the preferred embodiment, controller 170 is a powertrain controllerthat includes one or more digital microprocessors or digital computers,which cooperatively perform calculations, and execute subroutines andcontrol algorithms. Controller 170 generates a variable force solenoid(VFS) command or output signal, which controls the amount of slippagebetween the friction discs and spacer plates of clutch 100, therebycontrolling the relative magnitudes of torque and power transmitted tofront driveshaft 22 from the output 58 and rear driveshaft 26. The VFSsignal is the percentage of the cycle time for which the signal isactivated or enabled. The output signal of controller 170 iscommunicated to a variable force solenoid 256, which operates to openand close a source of fluid pressure 258 to a clutch servo 262, throughwhich clutch cylinder 96 is alternately pressurized and vented. Theclutch duty cycle is interchangeably referred to as a command, clutchoutput, and torque output.

In the preferred embodiment of the invention, memory 172 is a memoryunit including both permanent and temporary memory, and at least aportion of the operating software that controls operation of controller170. Moreover, memory 172 is adapted to selectively store other types ofdata or information, including data from prior control loop executions,processing data, and operational data. Examples of such data include,but are not limited to, data relating to the speed of driveshafts 18,20, the difference in speed of driveshafts 18, 20, referred to as “deltashaft speed” or clutch slip, and engine operating data, which are usedby controller 170 to determine the magnitude of torque currentlyproduced by the engine 169, and torque that should be transmitted byclutch 100. Processor 170 also accesses information it uses to determinethe current torque ratio of the torque converter 167 with reference tothe speed ratio across the torque converter, which is determined fromdata produced by speed sensors 168, 171. Memory 172 also stores variousfunctions, look-up tables, mathematical constants, threshold values, thestate of various flags, and the number in the register of variouscounters and timers.

Right and left front wheel speed sensors 174, 175, and right and leftrear wheel speed sensor 176, 177 are preferably speed sensors used in ananti-lock brake system (ABS). The speeds of the front sensors 174, 175are averaged, filtered and forwarded to controller 170 from an ABScontroller (not shown). Similarly, the speeds of the rear front sensors176, 177 are averaged, filtered and forwarded to controller 170 from the(ABS) controller. Controller 170 determines the speed of the frontdriveshaft 20 and speed of the rear driveshaft 18 from the processedoutput produced by the wheel speed sensors.

Electronic engine control unit 178 includes sensors for producingsignals representing engine operating parameters, which may include airmass flow rate, engine throttle position, barometric pressure, aircharge temperature, spark timing, engine speed NE, fuel flow, cam timingand other information used as indices to determine current engine outputtorque from data stored in computer memory. Generally, information fordetermining current engine torque is stored in the form of look-uptables or regressive polynomials established by dynamometer testing ofthe engine.

In the preferred embodiment, the EEC 178 includes a throttle positionsensor, which measures and/or detects the position of the enginethrottle, and transmits a signal representing this position tocontroller 170. EEC 178 may include one or more conventional engine,vehicle speed and/or acceleration sensors, and one or more faultdetection sensors, which detect faults or abnormalities in the operationof engine 169 or in the operation of the other components of vehicle.With reference to current engine operating condition data, EEC 178continually produces a signal representing current engine output torqueand throttle position, which signals are received as input by thecontroller 170. In addition, a foot brake switch 179 produces a high orlow signal representing the state of the foot brake. Alternatively, abrake position sensor continually produces an input signal representingthe current position of the foot brake pedal operated manually by thevehicle operator. The EEC 178 transmits the signal produced by sensor179 also to the controller 170.

Sensors 174-176 provide data representing measured operating parametersto controller 170, and they may include filtering and/or processingdevices or circuits, e.g., low-pass, high-pass, and/or band-passfilters, which filter and/or process the measured or sensed data priorto sending the data to controller 170.

Referring now to the strategy for controlling operation of clutch 100illustrated in the logic flow diagram of FIG. 4, at step 180 controller170 receives and processes the data representing shaft speed, anddetermines the speeds of the front driveshaft 20 and/or wheels 10 andthe rear driveshaft 18 and/or wheels 12.

At the position of a 4×4 switch, manually controlled by the vehicleoperator, is read to determine whether the switch is in the “auto”position, which indicates that clutch is controlled automatically by thestrategy of this invention. If the operator sets the switch to any otherof the selectable positions (4×2, 4×4 Low, and 4×4 High), the controlexits the clutch strategy at 184. When either 4×4 High or 4×4 Low areselected, clutch 100 is fully locked or engaged by pressuring the clutchservo sufficiently to engage the clutch. When the 4×2 range is selected,clutch 100 is open or disengaged by venting its servo. The status of the4×4 switch is tested during each execution of the control algorithm,each loop having a period of about loop 8 mS.

Provided the transmission range selector or PRNDL is in the D or Rrange, and if the inquiry at step 182 is true or high, control passes tostep 186, which produces a VFS signal that strokes clutch 100 for acalibratable period. Preferably line pressure normally in the range60-80 psi., is boosted is for about 150 mS, and the clutch pressure ismaintained thereafter at 20 psi., the steady state clutch pressuremagnitude or reference clutch pressure magnitude.

Reference to “calibrateable or calibrated” means a scalar or functionwhose value is a predetermined magnitude, which can be changed orcalibrated by altering the control algorithm to tune or produce adesired performance characteristic. Calibrated function values arestored in memory, the currently magnitudes of which are determined froma look-up table with reference to another variable or a set ofvariables, the arguments or indexes of the function.

Next at step 188, controller 170 calculates clutch slip or delta shaftspeed by subtracting the speed of the front driveshaft 20 (front shaftspeed) from the speed of the rear driveshaft 18 (rear shaft speed).

At step 190, the controller 170 determines whether the absolutedifference in speed of the front shaft speed and rear shaft is greaterthan a predetermined, reference slip magnitude. The reference magnitudeof slip is stored in a slip look-up table in memory 172. The look-uptable includes slip values, speed values, each of which corresponds to aparticular vehicle speed or range of vehicle speeds, and represents amagnitude of slip considered allowable at a particular vehicle speed orrange of vehicle speeds. The current vehicle speed is used to index orreference the look-up table, thereby providing a reference slip value.

If the current slip exceeds the maximum allowed or reference slip,indicating that slip is high, the control algorithm moves to a module192 of clutch control routines (slip mode, preemptive mode, closed loopmode, and handling mode), which are then executed.

If the current slip is equal to or less the maximum allowed slip orreference slip, control passes to step 194. A dual test is made at step194. First, the current throttle position (THROTTLE), a counted numberof clicks or another signal representing the extent to which thethrottle pedal is depressed, is compared to a coast throttle position(TP_COAST), a calibrated value, to determine whether THROTTLE is equalto or less than TP_COAST. A preferred value for TP_COAST is about 5-10%of full throttle. Second, the current clutch duty cycle (dc_cl) iscompared to a minimum clutch duty cycle value (DC_MIN), a calibratedvalue, to determine whether dc_cl is equal to or less than DC_MIN. Apreferred value for DC_MIN is about 2-5%. If both these tests are false,indicating both that the operator is demanding a high magnitude ofengine torque and the commanded clutch pressure is high, even though thevehicle is in a low slip condition, control passes to step 192. If step194 is true, control passes to step 196, where coast mode control isexecuted.

If a change in transmission gear ratio is in progress, represented bythe current presence of a flag indicating that a command has issued froma transmission controller for a ratio change and the ratio changes hasnot been completed, clutch pressure increases at a calibrated linearlyrate, holds at a calibrated duty cycle for a calibrated period, and thendecreases at a linear calibrated rate.

Upon entering module 192, at step 200 the counter, SLIP_CTR, isincremented after each execution loop of the control algorithm, therebykeeping an updated count of the number of execution loops of module 192,or length of time the control system 167 remains in module 192.

At step 202, controller 170 calculates an error signal ERR(k) bysubtracting Max_Allowable_Slip, a calibrated function of vehicle speed(VEH_SPD), from DELTA WHEEL SPEED (DELTA_WHEEL_SPD) or slip. Inperforming this calculation, the allowable slip factor is selected byway of a rear slip table, which includes a plurality of “allowable slip”values (e.g., speed values), each corresponding to a particular rearshaft speed value or range of values, and each representing an amount ofrelative slip that is considered to be allowable at a particular speedor range of speeds. The current rear shaft speed is used to index thetable, thereby providing a corresponding allowable slip value. In onenon-limiting embodiment, the control system 167 determines whether thewheels 10, 12 have differing effective diameters. If such a conditionexists, controller 170 increases the allowable slip value by an offsetfactor or value, thereby compensating for the differing effectivediameters of the wheels.

Controller 170 processes the wheel speeds produced by sensors 174-177and determines from those signals the speeds of the front driveshaft 20and rear driveshaft 18.

Preparatory to executing a closed loop control to determine a slipcontrol signal (dc_cl), proportional and integral (PI) constant valuesare determined. The PI signal y(k) is calculated by use of the followingequation:

y(k)=Y p(k)+Y i(k)  Eq. 1)

wherein k represents the current iteration of the calculation, Y i (k)is the integral term or component, and Y p (k) is the proportional termor component. The integral term is derived by use of the followingequation:

Y i(k)=Y i(k−1)+[T*K i]*ERR(k)  Eq. 2)

wherein T represents the time interval between iterations of thecalculation, K i is equal to an integral gain constant, and ERR(k).

The proportional term is calculated by use of the following equation:

Y p(k)=K p*ERR(k)  (Eq. 3)

wherein Kp is a proportional gain constant.

After the error signal ERR(k) is calculated at step 202, the controller170 proceeds to step 204 and selects a value for the proportional gainconstant K p. Values of K p are the product of a function stored inmemory indexed by current throttle position (THROTTLE) and currentvehicle speed (VEH_SPD), and a stored function indexed by ERR.

K p=FN(THROTTLE, VEH _(—) SPD)*FN(ERR(k))

In the preferred embodiment of the invention, K p is set or made equalto a “down” gain value “K p—DN” if e(k) is less than zero, whichindicates that the torque provided to the front driveshaft 20 should bereduced, and is set or made equal to a “up” gain value “K p—UP” if e(k)is greater than zero, which indicates the torque provided to frontdriveshaft 20 should be increased. By selectively using two separateproportional gain constant values is K p—UP and K p—DN, system 167 isable to substantially reduce noise, vibration and harshness withoutcompromising system response time.

Similarly at step 204, controller 170 determines a value for theintegral gain constant K i. Values of K i are the product of a functionstored in memory indexed by current throttle position (THROTTLE) andcurrent vehicle speed (VEH_SPD), and a stored function indexed by ERR.

K i=FN(THROTTLE, VEH _(—) SPD)*FN(ERR(k))

After the values of the integral and proportional constant terms K i andK p have been determined, at step 205 a constant K c is determined froma look-up table stored in memory containing values indexed byDELTA_WHEEL_SPEED, the difference in speed between the front and rearwheels 10,12 determine at step 188.

Kc is the overall gain of the PID controller, and is useful especiallyin calibrating the powertrain for increasing the effect of thecontroller without having to alter individual the values of theindividual PID gains.

Next, at step 208, the closed loop clutch duty cycle is calculated fromthe following equation and stored in memory:

dc _(—) cl=(ERR(k)*K p+ERR(k)*K i)*K c

A preemptive throttle position based clutch duty cycle (dc_pps) isdetermined next at step 210 with reference to the position of the enginethrottle, or the position of the throttle paddle for IR vehicles. Whenvehicle speed is high and the transmission is operating in a high gearratio, if the vehicle operator then tips into the throttle, i.e.,increases the extent to which the engine throttle is opened, a lowmagnitude of torque at the front wheels is required to prevent orpreempt wheel slip. However, when vehicle speed is low and thetransmission is operating in a low gear ratio, if the vehicle operatorthen tips into the throttle, a relatively high magnitude of torque atthe front wheels is required to prevent wheel slip. A first functionstored in memory, whose arguments or indices are vehicle speed andthrottle position, is established by calibration to control operation ofthe clutch 100 under the drive conditions described above. Accordingly,the values of that function, FN(VEH_SPD, THROTTLE), decrease withincreasing vehicle speed and increase with increasing throttle position.

However, it is undesirable to overload the front wheels 10 with hightorque for an extended period. Therefore, the preemptive clutch dutycycle determined with reference to the first function, FN(VEH_SPD,THROTTLE), is held only for a brief period, after which it is reduced toa steady state duty cycle, which is approximately zero percent. A secondfunction, FN(SLIP_CTR), is a shaping function for progressively reducingthe duty cycle generated by the first function. The value of SLIP_CTR, acount of the number of execution loops the control has been in the slipcontrol mode of module 192, is a measure of the length of the periodsince the control algorithm entered the slip control mode. The magnitudeof second function, indexed by the count within the current slipcounter, SLIP_CNT, varies inversely with the magnitude of the slipcounter. In this way, dc_pps declines as the duration of the slip modeincreases. During the occurrence of a slip condition, the control reliesprincipally on the closed loop control, to compensate for slip, ratherthan the throttle-based slip preemption control strategy.

The throttle position-based preemptive clutch duty cycle (dc_pps) isdetermined from the following equation:

dc _(—) pps=FN(VEH _(—) SPD, THROTTLE)*FN(SLIP _(—) CTR RT)

A preemptive throttle rate based clutch duty cycle (dc_pre) isdetermined next at step 212 with reference to the time rate of change ofthe engine throttle position, or, as it is called in connection with IRvehicles, the position of the throttle paddle. When throttle rate ishigh, either due to a throttle tip-in or back-out, i.e., movement of thethrottle tending to open or close the throttle, a relatively highmagnitude of torque at the front wheels is required to prevent orpreempt slip. However, when the throttle rate is low, a relatively lowmagnitude of torque at the front wheels is required to prevent slip dueto the inertia of the engine and transmission. At low vehicle speed withthe transmission operating in a low gear ratio, a relatively highmagnitude of torque at the front wheels is required to prevent slip. Aduty cycle function stored in memory, whose arguments or indices arevehicle speed and throttle rate, is established by calibration.Accordingly, the values of that function, FN(THROTTLE RATE, VEH_SPD),vary directly with throttle rate and inversely with vehicle speed.

However, it is difficult to rely on throttle position alone to controlslip following a tip-in condition. For example, if the throttle positionchanges from 0% to 20% in 100 mS, a 200% change of throttle position persecond, it has been discovered that it is preferred to maintain the dutycycle magnitude produced by the function FN(THROTTLE RATE, VEH_SPD) fora calibratable hold period of about 200-500 mS, and then decrease theduty cycle at a rate that decreases with time. The control strategy ofthis invention continually compares the current duty cycle to the baseduty cycle and continually decreases the current duty cycle at a ratethat increase with time. For example, after the hold period, dc_pre maybe reduced by 3% after the first execution loop, approximately 8 mS.Then it may be decreased by 2% after the second execution loop, and by1.5% after the third execution loop, etc. In this way, dc_pre quicklydeclines when dc_pre is substantially larger than the steady state orreference clutch duty cycle, and dc_pre declines more slowly as dc_preapproaches the steady state or reference clutch duty cycle.

The throttle rate-based preemptive clutch duty cycle (dc_pps) isdetermined from the following equation:

dc _(—) pre=FN(THROTTLE RATE,VEH _(—) SPD)*FN(SLIP _(—) CTR).

The throttle rate duty cycle is decremented or shaped on the basis ofthe following function:

FN(dc_cmd, dc_base)

Also located in the module 192 of control algorithms is a handling modeprogram entered at step 214 and executed in series with steps 200-212. Aprincipal purpose of the handling mode is to maintain clutch 100 in aslip condition, i.e., partially engaged, but not fully engaged. Aportion of the torque produced at the transmission output 38 istransmitted through the transfer case output 58 directly to the reardriveshaft 18 and rear wheels 12, and a portion of that torque istransmitted through clutch 100 and drive chain 90 to the frontdriveshaft 20 and front wheels 10. The relative magnitudes of thosetorque portions is determined by the degree of slip, the extent to whichclutch 100 is fully engaged.

The duty cycle produced by the control algorithm, dc_hndl, determinesthe magnitude of clutch slip during the handling mode. The handling modecontrol strategy decreases pressure to clutch 100 if the clutch islocked, and it simulates the functional characteristics of a centerdifferential and viscous coupling.

The handling mode control is illustrated in detail in FIG. 5. Tests aremade first at steps 216, 218, 220 to determine whether the handling modeshould be executed. The torque at the transfer case input 38 iscalculated as the product resulting from multiplying the current enginetorque, the current torque ratio of the torque converter, and thecurrent gear ratio of the transmission. Current engine torque isdetermined with reference to inferred engine output torque. The currenttorque ratio produced by the torque converter, which is located in thedriveline between the engine 169 and the transmission input shaft, isdetermined from the torque converter current speed ratio. The currentgear ratio of the transmission is calculated from the speed of thetransmission output shaft OSS 38 and speed of the transmission inputshaft, the torque converter turbine speed. Speed sensors 168, 171 andOSS produce signals representing the speed of the turbine, engine andtransmission output, respectively.

At step 216, a first test, used to determine whether the handling modeshould be executed, is a comparison of the difference between currenttorque at the transfer case input and initial input torque at thebeginning of the current execution loop. If that difference is greaterthan a predetermined threshold, a calibrated constant stored in computermemory 172, then the first test is true, and control passes to step 218.

A second test 218 compares the time rate of change of input torque to areset threshold, a calibrated constant. A third test 220 is determinedfrom the difference between the speed of the transfer case output shaft,which is directly connected to the rear drive shaft, and the speed ofthe front drive shaft. This speed difference represents the slip acrossthe clutch. If this speed difference exceeds another calibrated resetthreshold, then the third test is true. If either of the first, secondor third test is true, then the handling mode is entered because thevehicle drive line is in a dynamic state of torque change, to which thehandling mode is adapted.

A clutch hold timer is set equal to zero at 222 upon entering thehandling mode, whereupon the timer begins a count representing thelength of a period since the beginning of the current handling modeexecution. Continually during execution of the handling mode, the countwithin a clutch hold timer is compared at 224 to a threshold count, acalibrated reference value representative of a recent occurrence of adynamic state of torque change. If the clutch hold timer count is lessthan the threshold count, indicating that an insufficient amount of timehas passed since the last occurrence of a dynamic event, the handlingmode execution continues. Otherwise, control passes to step 238.

If at test 220, the clutch hold timer exceeds the reference time, whichwould indicates a dynamic change of torque has not occurred recently,the handling mode strategy concludes that the transfer case is operatingin a steady state mode. Consequently, the clutch duty cycle isrecalculated at 238 as the current duty cycle minus a calibrateddecrement rate. The clutch duty cycle is continually reduced in this wayuntil the clutch duty cycle becomes less than the steady state clutchduty cycle at test 240, whereupon the handling mode strategy is exitedat 230.

If the clutch hold timer is equal to zero, the test at step 226, theninitial input torque is set equal to the current input torque and thatvalue is saved at 228 in memory 172. If the clutch hold timer is equalto zero, indicating that the current execution loop is the first loop tobe executed since entering the handling mode, then at 228initial_InputTorque=input torque.

Next at 232 a target torque, representing the portion of the inputtorque that is to be transmitted from the transfer case output throughclutch 100 to the forward drive shaft 20, is determined with referenceto a predetermined, calibrated torque split scalar. Target torque isequal to the torque split scalar multiplied by input torque.

Then the handling mode clutch duty cycle is calculated at 234. First, alook-up function of linearized clutch duty cycle values, each valuecorresponding to the torque transmitted by clutch 100 at the duty cycle,is indexed by the current targetTorque value from step 232 to determinea clutch duty cycle. The value of this clutch duty cycle simulates theclutch torque that would be produced by a transfer case centerdifferential.

The second term of the clutch duty cycle summation simulates theperformance of a viscous coupling by increasing the magnitude of thehydraulic pressure sent to the clutch in proportion to the amount ofcurrent slip across the clutch 100. The proportional control componentis the product of a proportional constant and the magnitude of clutchslip, i.e., the difference between the speed of the rear drive shaft andthe speed of the front drive shaft.

Accordingly, either a constant, Proportional Constant, or a look-uptable of such values, is multiplied by the slip across clutch 100. Thedifference between the speed of the rear driveshaft 18 and the speed ofthe front driveshaft 20, (RearSpeed−FrontSpeed) is the slip across theclutch. The handling mode clutch duty cycle is calculated from thefollowing equation:

dc _(—) hndl=FN(target Torque)+Proportional Constant*(Rear Speed−FrontSpeed))

After the clutch duty cycle is calculated in this way, the clutch holdtimer is incremented at 236 each time the handling mode strategy isexecuted. The handling mode is repeatedly executed as long as theclutchHoldTimer count is less than the reference at step 224. When theclutchHoldTimer count exceed the reference count, control passes to step238, and the clutch pressure is reduced steadily until its correspondingclutch duty cycle magnitude becomes less than the steady state clutchduty cycle magnitude, whereupon the handling mode is exited. During thesteady state condition, the pressure at clutch 100 is no longerdecreased, but is held at a positive, low-pressure magnitude.

The handling mode strategy is reentered if any of the tests at 240 istrue. If the difference between input torque and initial input torqueexceeds an exit trigger Threshold, or the time rate of torque changeover a calibrated period is greater than an exit trigger Threshold, orclutch slip exceeds an exit trigger Threshold, then control passes tostep 214, and the handling mode is reactivated. Preferably, the exittrigger thresholds that lead to an exit of the handling mode strategyare somewhat higher that the reset thresholds used to enter the handlingmode.

After executing the slip mode, preemptive mode, handling mode and closedloop algorithms of module 192, as described, the controller 170determines the commanded clutch duty cycle by comparing at 250 the dutycycle values dc_pps, dc_hndl, dc_cl and dc_pre. The computer thenestablishes a commanded duty cycle signal, dc_cmd, the largest of theduty cycle values generated by executing the slip mode, preemptive mode,handling mode and closed loop mode algorithms. This commanded duty cycleis passed to step 252 for subsequent processing.

Returning again to FIG. 4, if the test at step 194 is true, indicatingthat the vehicle is at high speed, a low magnitude of engine torque isrequired, and the clutch duty cycle is at a low magnitude, the coast orengine braking mode algorithm is executed at 196. During a coastcondition, the transmission operates in a high gear ratio; therefore,engine braking is present. But engine braking applies a negative torqueto the rear wheels due to engine and transmission inertia. This negativetorque would slow the rear wheels relative to the front wheels tendingto destabilize the vehicle's rear. To avoid this condition, clutch 100is applied or engaged during an engine braking condition by a coast modecontrol algorithm.

An engine braking condition is detected when test 194 is true. A clutchduty cycle for the coast mode is determined at 196. A first function(FN(VEH_SPD)) containing calibrated clutch duty cycle values indexed byvehicle speed is established such that the first function's valuesincrease with vehicle speed. A second function (FN(BRK)) containingcalibrated values indexed by the state of the foot brake pedal isestablished such that the second function's values increase when thebrake pedal is displaced from a reference position. A third function(FN(SLIP_CTR)) containing calibrated values indexed by the value of theslip counter is established such that the third function's value changesinversely with the slip counter value. In this way, the magnitude of thefunction decreases steadily as the length of the clutch slip conditionperiod increases. Alternately, when a foot brake sensor of the type thatindicates the pedal displacement from a reference position is used, asecond function (FN(BRK)) containing calibrated values indexed by thedegree of foot brake pedal displacement is established such that thesecond function's values increases with brake pedal displacement.

The effect of the second function is to decrease the negative enginebraking torque applied to the rear wheels, and to slow the front wheelsby increasing the negative engine braking torque applied to the frontwheels. These actions are increased if and to the extent the vehicleoperator displaces the foot brake depending on the nature of the brakepedal sensor that is used. The effect of the third function is todecrease the magnitude of the coast mode clutch duty cycle steadily asthe length of the clutch slip condition period increases during anengine braking condition.

If output from a steering wheel angle sensor is available, then at 251the actual steering wheel angle is determined. Otherwise, the steeringwheel angle is inferred using the ISWA technique described in U.S. Pat.No. 6,498,975, which is incorporated herein by reference and is owned bythe assignee of this invention. The command value dc_swa is determinedfrom a stored function indexed by ISWA or actual steering wheel angle,if available, and vehicle speed. The commanded VFS magnitude is clippedto a maximum value in accordance with the value of dc_cmd, obtained froma stored function indexed by the current dc_cmd and dc_swa. The finaldc_cmd value is transmitted to the variable force solenoid 256 bycontroller 170.

At step 252, the controller 170 produces a final commanded duty cycle,dc_cmd, by comparing the duty cycle value generated at step 250following execution of the slip mode, preemptive mode, handling mode andclosed loop mode algorithms to the coast mode clutch duty cycle value.The computer then assigns to the commanded duty cycle, dc_cmd, thelarger of these clutch duty cycle values, and issues a commandrepresenting the final clutch duty cycle value.

The commanded duty cycle signal is carried as input to a signalconditioning circuit 254, whose corresponding output is a pulse widthmodulated VFS signal applied to a solenoid 256. A fluid pressure source258 is connected through a valve 260 controlled by solenoid 256. Valve260 opens and closes a connection between the fluid pressure source 258and a hydraulic servo 262 in response to the VFS signal. The pressure atservo determines the degree of engagement of clutch 100 by controllingthe force between the clutch spacer plates 88 and clutch friction discs86. In this way, the clutch slip is controlled in response to thecommanded clutch duty cycle signal (dc_cmd) produced by the controller170. The final commanded clutch duty cycle is the result of the severalduty cycle values resulting from the slip mode, preemptive mode, closedloop control, handling mode, and coast mode.

A technique is provided to protect clutch 100 against excessive heatproduced by dissipating energy while changing the duty cycle of theclutch in response to the commanded duty cycle. The clutch heatprotection mode is executed at step 270. The heat protection controlcontinuously determines the operating temperature of the transfer case.The control disables the handling mode 214 and directs control to thepreemptive traction mode 210, 212 when a calculated clutch temperatureexceeds a first threshold temperature. Alternatively, if the temperatureof the calculated transfer case produced by the control is greater thana second threshold, the heat protection control commands an auto-lockmode, which fully locks or engages the transfer case clutch 100, causingthe transfer case to operate in the four wheel drive high range, 4×4High.

The output torque of the transmission is equal to the transfer caseinput torque due to the direct connection of the transmission outputshaft 38 to the transfer case input, as shown in FIG. 2. Engine outputtorque is inferred at step 272 by the EEC 178 from the engine operatingparameters and either a regressive polynomial or a function of enginetorque stored in the EEC 178 and indexed by the current engine operatingparameters, as described above.

At 274 the controller 170 calculates transfer case input torque usingengine output torque, the gear ratio of the transmission, and currenttorque ratio produced by the torque converter 167. The torque ratioproduced by the torque converter is determined directly from the currentspeed ratio produced by the torque converter, the ratio of the torqueconverter turbine speed TSS. Speed sensor 168 produces a signalrepresenting the torque converter turbine speed, and speed sensor 171produces a signal representing engine speed.

Next, at step 276 the controller calculates the magnitude of power inputat clutch 200, which is the product of the transfer case input torqueand the speed of the input element 82 of the transfer case clutch, i.e.,the speed of output 58. Alternatively the controller calculates thespeed of the input element 82 of the transfer case clutch 100 from theturbine speed sensor 168 signal, the current transmission gear ratio andthe current gear ratio produced by the transfer case gearset 44.

Then, at 278 the output torque at the transfer case clutch 100 iscalculated as the product of the hydraulic pressure at the clutch andthe clutch gain. The magnitude of pressure in the clutch servo 262 is aninput to the controller from a pressure sensor at the servo 262 or isknown from the current commanded duty cycle (dc_cmd) and the pressure ofthe pressure source. The controller memory 172 contains the values ofclutch gain, which is the product of the average coefficient of frictionof a clutch friction disc 88-spacer plate 86 pair, the number of suchpairs of discs and plates, the effective friction area of a pair ofdiscs and plates, and the effective radius of the frictional area fromthe clutch axis of rotation. Generally, the clutch gain is in the range2-5 foot-pounds/psi. The clutch servo fluid pressure is known to thecontroller 170 from the current commanded clutch duty cycle or ameasured input to the controller.

Next, the controller 170 calculates at step 280 the power output by thetransfer case clutch 100. Clutch output power is the product of theoutput torque at the clutch and the speed of the clutch output, which isthe speed of the output shaft 58 minus the slip across clutch 100.

The controller calculates at step 282 the heat power dissipated byclutch 100 as the difference between the transfer case clutch inputpower and the clutch output power. Then the controller integrates theheat power dissipated by clutch 100 over the period of each executionloop, approximately 8 milliseconds, to determine the heat energy inputto the transfer case clutch during the current execution loop. Thethermal mass of the clutch 100 is a known constant stored in memory 172.

The controller divides the heat energy input to the clutch during thecurrent execution loop by the thermal mass of the clutch to determine atstep 284 the change of clutch temperature during the current executionloop. The controller maintains a running updated sum of clutchtemperature derived from these incremental clutch temperature changes.At step 286, the temperature during the next execution loop is set equalto the current clutch temperature plus the incremental temperature ofthe current loop.

The clutch temperature, calculated in this way, is compared at 288 to afirst, Preemptive Mode calibrated threshold temperature. If the test at288 is true, control returns to step 272 and the steps are executedagain during the next loop.

If the clutch temperature at 286 exceeds the first thresholdtemperature, an inquiry is made at 290 to determine whether clutchtemperature is equal to or greater than an Auto Lock referencetemperature, which is higher than the first threshold temperature. Ifthe test at 290 is true, the controller produces a commanded duty cyclesignal to solenoid 256, which supercedes the handling mode duty cycle,dc_hndl, increases the pressure at servo 262, and fully engages clutch100 in the auto lock mode, wherein the transfer case operates in the 4×4High range. If the test at 290 is false, at 294 the controller producesa command to enter the preemptive handling mode, causing the tractionmode commanded duty cycle to control operation of clutch 100.

Although the form of the invention shown and described here constitutesthe preferred embodiment of the invention, it is not intended toillustrate all possible forms of the invention. Words used here arewords of description rather than of limitation. Various changes in theform of the invention may be made without departing from the spirit andscope of the invention as disclosed.

We claim:
 1. In a vehicle drive line having an engine controlled by athrottle position, a method for controlling a clutch that driveablyconnects an input and an output with varying degrees of clutchengagement, the method comprising the steps of: operating the clutchpartially engaged; determining a current throttle position, throttlerate, and vehicle speed; determining a first desired clutch engagementcorresponding to the current throttle position and vehicle speed;determining a second desired clutch engagement corresponding to thecurrent throttle rate and vehicle speed; and changing the degree ofclutch engagement to the greater of the first desired clutch engagementand the second desired clutch engagement.
 2. The method of claim 1,further comprising: determining a length of a first period that beginsupon initiating partial engagement of the clutch; reducing the firstdesired clutch engagement by a factor whose magnitude varies inverselywith the length of the first period; and changing the degree of clutchengagement to the greater of the first desired clutch engagement and thesecond desired clutch engagement.
 3. The method of claim 2, furthercomprising: maintaining the first desired clutch engagement for a secondperiod of predetermined length.
 4. The method of claim 1, furthercomprising: determining a reference clutch engagement; determining adifference between the second desired clutch engagement and thereference clutch engagement; reducing the second desired clutchengagement by a magnitude that varies directly with said difference; andchanging the degree of clutch engagement to the greater of the firstdesired clutch engagement and second desired clutch engagement.
 5. Themethod of claim 1, wherein the step of reducing the second desiredclutch engagement further comprises: subtracting from the second desiredclutch engagement a magnitude that decreases as the magnitude of saiddifference decreases.
 6. The method of claim 1, wherein the step ofoperating the clutch in a partially engaged condition includes the stepsof: determining a current clutch slip; establishing a first desiredportion of the input torque to be transmitted by the clutch to thesecond output; determining a first magnitude of clutch torquecorresponding to the first desired portion; determining a secondmagnitude of clutch torque to be transmitted to the second output inproportion to the current clutch slip; and changing the magnitude oftorque transmitted by the clutch to the sum of the first and secondmagnitudes.
 7. The method of claim 1, wherein the step of increasing thedegree of clutch engagement over a period sufficient to reduce thecalculated temperature of the clutch includes the step of fully engagingthe clutch.
 8. A method for controlling, with the aid of a digitalcomputer, a clutch that driveably connects an input and an output withvarying degrees of clutch engagement, the clutch operating in a vehicledriveline that includes an engine controlled by throttle position, themethod comprising the steps of: inputting to and executing in thecomputer a computer readable program code algorithm for operating theclutch partially engaged; providing the computer with a data baseincluding at least a first desired clutch engagement that varies withthe current throttle position and vehicle speed, and a second desiredclutch engagement that varies with a current throttle rate and a currentvehicle speed; providing a signal in the computer representing thecurrent throttle position, and vehicle speed; repetitively determiningin the computer at frequent intervals during execution of the algorithmthe current throttle rate, the first desired clutch engagementcorresponding to the current throttle position and vehicle speed, andthe second desired clutch engagement corresponding to the currentthrottle rate and vehicle speed; issuing from the computer a commandclutch duty cycle signal representing the greater of the first desiredclutch engagement and the second desired clutch engagement; and changingthe degree of clutch engagement in response to the command signal. 9.The method of claim 8, wherein the step of issuing a command from thecomputer further comprises: terminating execution in the computer of thealgorithm; and issuing from the computer a command clutch duty cyclesignal causing the degree of clutch engagement to change to full clutchengagement in response to the command signal.
 10. The method of claim 8,further comprising the step of: initiating a counter in said computerupon the beginning of execution of the algorithm for monitoring thenumber of executions by the computer of the algorithm; providing thecomputer with a data base that further includes a factor whose magnitudevaries inversely with a current count; determining from the computerdata base the magnitude of the factor that corresponds to the currentcount of the counter; repetitively reducing in the computer the firstdesired clutch engagement by the magnitude of the factor; and issuingfrom the computer a command clutch duty cycle signal representing thegreater of the first desired clutch engagement and the second desiredclutch engagement, whereby the degree of clutch engagement is changed inresponse to the command signal.
 11. The method of claim 8, furthercomprising the step of: providing the computer with a data base thatfurther includes a reference clutch engagement; repetitively determiningin the computer at frequent intervals during execution of the algorithma difference between the second desired clutch engagement and thereference clutch engagement; repetitively reducing in the computer atfrequent intervals during execution of the algorithm the second desiredclutch engagement by a magnitude that varies directly with saiddifference; and issuing from the computer a command clutch duty cyclesignal representing the greater of the first desired clutch engagementand the second desired clutch engagement, whereby the degree of clutchengagement is changed in response to the command signal.
 12. The methodof claim 11, wherein the step of reducing the second desired clutchengagement further comprises: during each execution of the algorithm bythe computer, subtracting from the second desired clutch engagement amagnitude that decreases as the magnitude of said difference decreases.13. The method of claim 11, further comprising: providing the computerwith a data base that further includes a reference period length;continuing to issue from the computer for the reference period length acommand clutch duty cycle signal representing the greater of the firstdesired clutch engagement and the second desired clutch engagement,whereby the degree of clutch engagement is changed in response to thecommand signal.
 14. In a transfer case, driveably connected to an enginecontrolled by throttle, a system for controlling a clutch that driveablyconnects the first output and second output with varying degrees ofclutch engagement, comprising: means for operating the clutch partiallyengaged; means for determining a current throttle position, throttlerate and vehicle speed; means for determining a first desired clutchengagement corresponding to the current throttle position and thecurrent vehicle speed; means for determining a second desired clutchengagement corresponding to the current throttle rate and the currentvehicle speed; means for producing a command clutch duty cycle signalrepresenting the greater of the first desired clutch engagement and thesecond desired clutch engagement, whereby the degree of clutchengagement is changed in response to the command signal.
 15. The systemof claim 14, further comprising: a fluid pressure source; a servothrough which the clutch is pressurized from the pressure source tochange the degree of clutch engagement; and a valve operated by asolenoid for opening communication between the pressure source and theservo in response to the clutch duty cycle command signal applied to thesolenoid.
 16. The system of claim 14, further comprising: means fordetermining a length of a first period that begins upon initiatingpartial engagement of the clutch; means for the first desired clutchengagement by a factor whose magnitude varies inversely with the lengthof the first period; and means for changing the degree of clutchengagement to the greater of the first desired clutch engagement and thesecond desired clutch engagement.
 17. The system of claim 14, furthercomprising: means for maintaining the first desired clutch engagementfor a second period of predetermined length.
 18. The method of claim 14,further comprising: means for determining a reference clutch engagement;means for determining a difference between the second desired clutchengagement and the reference clutch engagement; means for reducing thesecond desired clutch engagement by a magnitude that varies directlywith said difference; and means for changing the degree of clutchengagement to the greater of the first desired clutch engagement andsecond desired clutch engagement.
 19. The method of claim 14, whereinthe step of reducing the second desired clutch engagement furthercomprises: subtracting from the second desired clutch engagement amagnitude that decreases as the magnitude of said difference decreases.